Resonant piston pumps

ABSTRACT

A RESONANT PISTON PUMP WHEREIN SPRING MEANS ARE PROVIDED WHICH ARE SECURED AT ONE END TO THE PUMP HOUSING AND AT THE OTHER END TO THE PUMP PISTON AND BEING OPERATIVELY ARRANGED AND ASSOCIATED WITH SUCH PISTON AND HOUSING TO PROVIDE FOR SUBSTANTIALLY PARALLEL MOVEMENT OF THE PISTON WITHIN BUT OUT OF CONTACT WITH THE PUMP CYLINDER DURING ALTERNATE COMPRESSION AND EXPANSION OF THE SPRING MEANS. THE PISTON IS DRIVEN BY AN ELECTROMAGNETIC DRIVE MEANS IN A RESONANT MODE OF VIBRATION WITH THE SPRING MEANS TO PROVIDE FOR MAXIMUM POWER TRANSFER AND EFFICIENCY OF PUMP.

States Patent Inventors Peter W. Cull-wen Ballston Spa; Richard V.Newell, Loudonville, NY. Appl. No. 882,446 Filed Dec. 5, 11969 PatentedJune 28, 1971 Assignee Mechanical Technology incorporated Latham, NY.

RESONANT PISTON PUMPS 114 C 16 Drawing Figs.

1.1.5. (11 417/417, 417/571 int. Cl ..F04b 35/04, F04b 21/02 Field MSearch 417/416, 417, 363, 328, 418, 371

[56] References Cited UNITED STATES PATENTS 2,721,024 10/1955 Zeh417/417 3,325,085 6/1967 Gaus 417/416 Primary Examiner-Robert M. WalkerAttorney-Joseph V. Claeys ABSTRACT: A resonant piston pump whereinspring means are provided which are secured at one end to the pumphousing and at the other end to the pump piston and being operativelyarranged and associated with such piston and housing to provide forsubstantially parallel movement of the piston within but out of contactwith the pump cylinder during alternate compression and expansion of thespring means. The piston is driven by an electromagnetic drive means ina resonant mode of vibration with the spring means to provide formaximum power transfer and efficiency of the pump.

PATENTEU JUH28 m1 3,588,291

SHEET 1 BF 5 fm/end-or's eter Garwew fi /chard We el/PATENTEDJUP-i28l97l 3,588,291

SHEET 2 UF 5 [M Mentors We e r Cwuugw Q/W Wowwwy PATENTEDJUHZEHBH3588.291

SHEET 3 OF 5 I? van toms et-er" Curwen f/C/v'awa The invention describedherein relates generally to pumps and more particularly to an improvedelectromagneticallydriven resonant piston pump having a broad range ofapplica tions including aerospace, industrial and commercial.

The term pump" is used herein in the broad sense and includes devices ormachines that raise, transfer, compress or other wise operate on fluidsor attenuate gases. Although, as indicated, the invention has a widerange of applications, it is especially useful in connection withcompressors and will be particularly described in that connection.

Many different designs of resonant piston pumps have been proposed foroperating with either a liquid or gaseous medium. These designs normallyinclude a cylinder having intake and discharge valves, one or morepistons in the cylinder, and a spring associated with the piston toprovide a spring-mass system. To obtain piston movement for pumping afluid, a solenoid equipped with a movable iron armature is connectedwith the piston, so that when the solenoid is energized, it produces anelectromagnetic force which engages the armature and drives thespring-mass system at the same frequency as the varying electromagneticfield. By designing the resonant frequency of the pump spring masssystem to be essentially the same as the frequency of theelectromagnetic field, the pump- ,ing work can be performed at highefficiency.

Resonant pumps of the prior art utilizing the above major componentsgenerally have not found acceptance in the industrial community,primarily because both the piston and cylinder walls experienceexcessive wear within a time much shorter than that considered normalfor other types of reciprocating pumps. in general, such wear resultsfrom the combined effects of a number of independently acting forceswhich bear on the moving parts. For example, springs or other devicesassociated with the piston exert forces which fail to keep the pistoncentered on the cylinder axis during operation. Although the prior artdiscloses special designs of piston centering devices for use inresonant pumps, none of such disclosures have resolved the problem withany reasonable degree of success. Also, an inability of the solenoid togenerate a uniform flux field or to at least cause the armature toalways move on the cylinder axis contributes importantly topiston-cylinder wear. Liquid lubricants and piston rings made ofdifferent kinds of materials having self-contained lubri cants, alsohave been employed in minimizing the piston and cylinder wear problem inan attempt to obtain long operating life, but have met with onlymarginal success. Moreover, such pumps are not suitable for use inenvironments requiring contaminant-free atmospheres or where the processfluid being pumped must remain free of foreign particles.

To resolve the above and other problems, a major development programaimed at designing and manufacturing an efficient, reliable, gaslubricated resonant pump was undertaken by applicants assignee, with theresult that the prior art problems were successfully overcome andoutstanding pump performance has been achieved. The completely newdesign of pump developed is described and claimed US. Pat. Nos.3,156,405, 3,303,990 and 3,329,334, assigned to the same assignee as thepresent invention. The major improvements included the development of anew type spring of U shape configuration and in designing the pistonouter surface to provide for hydrodynamic or hydrostatic gas lubricationof the piston during its reciprocating movements in the compressor, thuscompletely eliminating the need for oil or other kinds of lubricants.The U-shaped springs were spaced at 90 intervals with one end connectedto the pump housing and the other to the piston so that upon pumpoperation, the action of the gas and electromagnetic forces acting onthe piston, coupled with the spring forces, causes the piston to move onthe cylinder axis, thereby eliminating problems of wear caused byrubbing ofthc parts as in prior art constructions.

Since potential applications were envisioned where long life is anecessity, gas bearing technology concepts were applied to therelatively moving parts, thus eliminating the need for solid or liquidlubrication of the moving parts as required by prior art designs. Asillustrated in the aforementioned patents, the only moving parts are asolenoid armature, springs, and valve means. Still fewer moving partsare required when the piston inlet and outlet ports are used instead ofvalve means. Evidence of success of the design resides in the fact thatthe pressure and flow rates have been maintained at the design pointafter several thousand hours of testing and trouble-free operation.

Although the new and improved compressors described and claimed in theforegoing referenced patents have been, completely successful, thereremains a continuing need for further research and development toprovide still more improved high performance, lubricant-free pumps whichare not only less expensive and bulky but also exhibit the same orbetter reliability and efficiency.

Applicants prior resonant pumps have always included design features forlubricating the relatively moving parts with a gaseous medium. Doing sorequired that substantial consideration be given to the pump designduring its early stages to assure having a design of piston, cylinderand springs such that hydrostatic or hydrodynamic lubrication of thepiston would result when the pump was placed in operation. Applicantsdevelopmental efforts now show that with the correct design andselection of parts, the prior art lubrication problems may beessentially eliminated.

SUMMARY OF THE INVENTION A primary object of the invention, therefore,is to provide an improved resonant piston pump which eliminates the needfor lubricants of any kind between the relatively moving parts bymaintaining the piston centered in the cylinder during all modes ofoperation.

Another object of the invention is to provide a resonant spring whichwhen installed in the pump, minimizes the lateral and bending vibrationswhich heretofore caused the piston to contact the cylinder walls.

Still another important object of the invention is the provision of ahighly reliable resonant piston pump of simple, com pact and economicalconstruction.

In carrying out the invention, the need for any lubricant in a resonantpump is obviated by choosing helical springs of the correct material andcharacteristics suitable for certain sizes of pump, and positioning suchsprings in a predetermined pattern around the piston. By then connectingthe spring ends, respectively to the housing and piston, the lateral andbending vibrations of the spring heretofore encountered are reduced tothe point where the piston is caused to move substantially parallel withthe cylinder walls thus minimizing contact between the piston andcylinder, and reducing piston-cylinder wear to negligible amounts. Bydesigning the springs to coact with the piston in a particular manner,the size of the pump is substantially reduced while obtaining optimumperformance through the use of conventional port or valve means andother pump components. It readily will occur to those skilled in the artthat the teachings of this invention are applicable to designs ofmachines or equipment other than resonant compressors and that thelatter is used only to illustrate the preferred embodiments of theinvention.

BRIEF DESCRIPTION OF THE DRAWINGS While the specification concludes withclaims particularly pointing out and distinctly claiming the subjectmatter of the invention, it is believed the invention will be betterunderstood from the following description taken in connection with theaccompanying drawings wherein:

FIG. 1 is a sectional view in elevation ofonc embodiment of a resonantpiston compressor designed in accordance with the teachings of thisinvention.

FIG. 2 is an end view of the compressor of FIG. 1 with parts broken awayto show the arrangement of discharge valves.

FIG. 3 is a perspective view of only the compressor housing illustratingthe inlet and outlet passages for the introduction and discharge of gasfrom the compressor.

FIG. 4 is a plan view of a valve having flexible fingers for controllingthe gas discharge passages in the compressor.

FIG. 5 is a sectional view in elevation of a modification of theinvention showing a somewhat different arrangement of valves andresonant spring used in the compressor;

FIG. 6 is a plan view ofthe compressor shown in FIG. 5.

FIG. 7 is a view taken on lines 7-7 of FIG. 6 showing the compressor ina different angle of orientation.

FIG. 8 is a view taken on lines 8-8 of FIG. 7 showing the disposition ofsprings in the compressor.

FIG. 9 is a view taken on lines 9-9 of FIG. 5.

FIG. 10 is a perspective view of the valving arrangement shown in FIGS.5-9.

FIG. 11 is an enlarged section view taken on lines 11-11 of FIG. 9.

FIG. 12 is a view in elevation, partly in section, of a modificationshowing a different arrangement of resonant springs used in thespring-mass system of the compressor.

FIG. 13 is a view taken on lines 13-13 of FIG. 12.

FIG. 14 is a view in elevation, partly in section, of another embodimentof the invention.

FIG. 15 is a view taken on the right side of the FIG. 14.

FIG. 16 is a section view of the compressor shown in FIG.

14 taken along line 16-16.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS Referring now to thedrawings wherein like reference characters designate like orcorresponding parts throughout the several views, there is shown inFIGS. 1-4, a resonant compressor in accordance with one embodiment ofthe invention including a housing 20 having a cylinder 22 and a cylinderhead 24 designed for receiving a discharge valve 26. A hollow piston 28having skirts 30 is adapted for reciprocation within the cylinder, and amachined helical spring 32 designed to a size and configuration to fitwithin the piston forms part of a spring-mass system as fully describedhereafter. The magnetic forces necessary to effect piston movement aregenerated by a solenoid 34 attached to the other end of the housing,Solenoid 34 comprises a conventional coil 36 wound ona laminated ironcore 38 held together by bolts or other fastening means. Anarmature 40designed for coaction with the solenoid, is connected to the piston 28by an interconnecting rod or bolt 42. One end of the bolt 42 is attachedto the armature at 44 while the other end extends through a holecentrally bored in the piston head and is secured thereto by a nut 46.As shown in FIG. I, the cylinder head cavity 47 is designed to receiveand accommodate the projecting nut 46 when the piston reaches the top ofits stroke.

Although the material composition of the piston and cylinder usually aredifferent, they nevertheless should have compatible characteristics,such as high dimensional stability, similar coefficients of thermalexpansion, good rubbing characteristics and be corrosion resistant. Thepiston should preferably be made of a material having low mass densityso as to minimize stiffness (and thus vibratory force and alternatingstress levels) of the resonant spring. In the particular design shown,the piston was made of a carbonaceous material while the cylinderconsisted ofa high-nickel content cast iron.

The specific bending and lateral vibration problems encountered in theprior art resulting from utilizing a helicallyshaped spring forachieving resonance, arose out of the use of a spring having amultiplicity of turns with one end attached to or supported by astationary member and the other end connected to or otherwise associatedwith the piston. Depending on the particular compressor design, as thespring was compressed and extended during the compression and expansionstrokes, the resulting spring forces produced a moment on the pistonwhich caused the piston to simultaneously rotate about a transverse axisand to laterally deflect. This caused contact and consequent wear of thepiston and cylinder wall surfaces during operation.

To overcome this problem in the present invention, a helical spring isprovided having upper and lower end plates with three or more separatesprings interposed therebetween with the coils of each spring lying onthe cylinder axis and in the same cylindrical plane, thereby achievingsymmetry in the structure. The complete spring assembly was machinedfrom solid stock such that the opposite ends of each spring wereintegrally formed with upper and lower plates respectively and the endswere spaced 120 from each other on each plate. Since these three coilsare symmetrically spaced and located between the end plates, uponapplication of a compressive or expansive force the spring forces willbe axis symmetrically distributed on the end plates, and rotation orlateral deflection of the piston in the cylinder therefore will not takeplace.

Referring again to FIG. 1, the spring 32 includes three separate andparallel spring coils 50, 51 and 52 interposed between the end plates 48and 49 to form the integral spring assembly. The ends of the threeseparate coils merge into their respective end plates 48 and 49 atpoints spaced 120 apart to assure that the resultant of the three springforces is transmitted through the spring-assembly centerline, therebyassuring that the piston 28 will move parallel to the cylinder walls. Toprovide both an efficient and'practical means for clamping the springbody in position, the plate 49 is equipped with an integrally formedflange 53 which extends radially outward from the plate 49; the flangebeing adapted for positioning between the end of the housing 20 and asupport member 54 to which the iron core 38 is attached.

By utilizing this kind of arrangement having three or more symmetricallyplaced spring coils with a corresponding number of ends merged into thespring end plates, the unequal moments resulting from the spring forcesare balanced out, thus eliminating the tendency of the spring, andtherefore the piston, to rotate or tilt and to move laterally in thecylinder. Although torsional forces tending to rotate the piston aslight amount about the centerline of the cylinder are still present,they do not effect lateral movement or transverse rotation of the pistonand contact between the piston and cylinder walls does not occur.

To provide for intake and discharge of gas from the compressor, inletports 56 are circumferentially disposed around the cylinder, and anoutlet 58, FIG. 3, which communicates with the discharge manifold 70,exhausts the compressed gas to the system. As shown more clearly inFIGS. 1, 2 and 4, the valve consists of a flexible plate 60 having anumber of cantilever reeds 62 which are elastically deflected to an openposition to uncover discharge openings 64 in the cylinder head when thecompression chamber pressure exceeds the pressure in the dischargemanifold 70. One end 66 of the discharge valve is anchored in place by aplate 68 secured to the cylinder head 24 by bolts 69, while the otherend terminates in projecting reeds 62 which are free for movement in adirection outward from the cylinder head to the open position.Compressed gas flowing from the compression area into the openings 64and passed the valve, discharges into a manifold 70 and outlet 58 priorto delivery to an attached system.

In unusual applications where long life and maintenancefree operation isa necessity, to further assure that the piston will not rub or otherwisecontact the cylinder walls, a multiplicity of gas inlet ports 72 may beincorporated in the cylinder walls for delivery of a gas forhydrostatically lubricating the piston and cylinder walls. Although thegas lubricant may be supplied from a separate source, the embodimentillustrated in FIGS. 1-3 shows an arrangement wherein the lubricant isfurnished by the compressor itself. Fluid is taken from manifold 70through outlet 74 to passages 76 which communicatc with the gaslubricant inlet ports 72. Should the piston be displaced off thecylinder centerline for any reason, a consequent increase in fluidpressure will occur in the areas of least clearance. Likewise, thepressure in the piston-cylinder clearance space on the other side of thepiston will decrease. As a result of the differential pressures, therelatively higher pressure fluid on'one side will act on the exposedpiston surfaces and keep the piston centered.

The compressor may be made to operate from either a direct current oralternating current source. In the embodiment illustrated herein, adiode is connected with the solenoid coil to furnish a pulsating currentfor driving the armature and connected piston at a 60 Hz. frequencyalthough such diode is not required in all cases. The required amplitudeof the compressor-frequency harmonic driving force can be calculatedclosely according to conventional design practices. Also, the mechanicalcomponents comprising the piston 28, spring 32, armature M) and shaft 42connecting the armature and piston, constitute a mass-flexure systemwhich is driven electromag netically and designed to resonate at 60cycles per second. To show relative values, in one compressor design,the total compression power at p.s.i.a. inlet pressure was calculated tobe 50 watts. For a 0.42 inch piston stroke and a 60 Hz. compressorfrequency, the resulting 60 Hz. harmonic force amplitude was 11.0 lb. Itwill be evident that the mass-flexure system must be made compatiblewith the electromagnetic driving force to secure a resonant mode ofoperation to obtain the generation of maximum power and efficiency inthe system.

In operation, upon energizing the solenoid with a suitable AC voltage,the pulsed current flowing in the solenoid coil produces flux whichlinks the armature. An electromagnetic force is thus produced whichdraws the armature into the iron core of the solenoid. Simultaneous withthat action, the spring 32 compresses and reaches the full degree ofcompression when the armature completes the end of its travel into thecore. As the electromagnetic forces commence moving the armature out ofthe core, the spring starts to expand and the piston moves to coverports 56 and start compressing the gas trapped in the cylinder.Continued expansion of the spring with consequent movement of the pistoncompresses the gas in the cylinder until its pressure exceeds that ofmanifold 70, whereupon the reeds 62 of valve 26 are displaced from theirseat and gas from the cylinder is then discharged into the manifold fordelivery to the attached system.

During the compression stroke, the piston does not contact the cylinderwalls because the separate spring coils 50, 51, and 52 integrally formedwith the plates 48 and 49, uniformly exert equal and symmetric forces onthe piston causing it to move along the cylinder axis. The resultantforce thus exerted on the piston passes directly through its axis andthe piston therefore does not tilt in the cylinder. When the gas incompression chamber 22 equals the pressure in manifold 70, the valve 26closes and the spring 32 which at that time is in an expanded condition,starts moving the piston on the return stroke (to the right as indicatedin FIG. 1). At this time, the electromagnetic forces generated by thesolenoid coil 36 assists in drawing the armature and the attached pistoninto its opening in the solenoid core. Continued piston movementuncovers ports 56 to permit the entry of a fresh supply of gas forcompression and the cycle is then repeated. During piston movement, gasfrom the discharge of the compressor, or from a separate source, may besupplied to inlets 72 to furnish a lubrication medium to thepiston-cylinder clearance space to help assure maintaining the piston onthe cylinder axis in the manner described above.

As indicated previously, a mass-flexure system comprising the spring,piston and armature is chosen to have a natural vibration frequencyessentially the same as the frequency of electromagnetic flux generatedby the solenoid coil. As these two frequencies coincide duringoperation, a resonant mode of vibration is established which producesboth maximum power and efficiency in the compressor. Successfulperformance of the compressor manufactured in accordance with the abovedesign has been achieved and such success is primarily attributable tothe use ofa multiple coil spring of the design illustrated in thedrawings. Even better performance has been achieved by utilizing a gasfor lubricating the cylinder walls in the piston-cylinder clearancespace since the gas bearing action produced by the relatively movingparts causes the piston to act like a bearing and stay on the cylinderaxis while simultaneously producing friction-free operation. Sincereliability and efilciency were important factors considered in thedesign, a combination of inlet ports and discharge valves were used,rather than full porting arrangements, although it is obvious thateither one or a combination can be used.

FIGS. 5-11 illustrate another embodiment of compressor in accordancewith the invention although the basic components remain the same. Theimportant changes include decreasing the distance between the armatureand piston to obtain better stability-and in locating the spring outsidethe cylinder walls. This minimizes the moments which resultkfromunbalanced lateral electromagnetic forces which act on the lateral polefaces at the solenoid armature. By decreasing this distance, thecompressor can also be packaged into a more compact construction sinceit permits locating the resonant spring outside the cylinder and reducesthe overall length of the machine. Other changes will be apparent as thedescription proceeds, and referring more specifically to the drawings,the solenoid 34 comprises a coil 36 wound on a laminated iron core 38,and an armature 40. A plate has one end secured to the armature 40 whilethe other end terminates in an enlarged head 82. A bolt 84 having itshead 86 fitted in a depression in the piston surface interconnects thepiston 20 with the armature, the arrangement being such thatelectromagnetic forces acting on the armature move the piston axially toobtain compression of a gas in the cylinder. The resonant spring 32positioned outside the cylinder is arranged for compression andexpansion as the piston moves in the cylinder. To obtain piston movementin a line parallel with the cylinder walls, the foreces tending tocompress the spring reaction forces respectively are uniformly appliedto the housing and the piston. The uniform application of force isobtained by utilizing three separate spring coils 87, 88 and 89 nestedtogether to form the spring body 32, with their ends attached to animmovable supporting structure 90. The spring ends are circumferentiallydisposed from each other at l20 intervals. As shown, a

bracket 90 attached to the housing includes a clamp 92 and.

bolt 93 for firmly grasping and holding each spring coil in a presetposition. The other ends of the nested spring coils are supported in amovable yoke 94 which comprises a hub 96 held in place between thepiston 20 and armature 40 by bolt 84. Three yoke arms project radiallyoutward from the hub and respectively terminate in cylindrically shapedmembers 95' which serve to anchor the other ends of the spring coils inposition. Cylindrically shaped members 95 may also have a clamp and boltfor firmly grasping and holding each spring coil.

It will be apparent that the yoke effectively serves as a connectionmeans between the spring coils, piston and the solenoid armature.Because of the symmetrical arrangement of the spring coils 87, 88 and89, and the location of brackets 90 on the housing and member 95 on theyoke for holding the spring ends, the forces applied to the spring bodyduring the piston compression stroke, are uniformly transmitted toequally spaced points on the housing, thus causing the piston to move onthe cylinder axis. Likewise, the piston is not diverted from movement onthe axis during the return stroke by the expanding springs because thespring reaction forces are absorbed uniformly by the spring supportingstructure.

The cylinder head 243 closes the upper end of the compression chamberand serves as a base for valves controlling the inlet and discharge ofgas to and from the compression chamber. The head is attached to thehousing by bolts R00 and is equipped with an inlet I02 and na outlet104. The inlet preferably is equipped with a filter I06 for removingforeign particles from the gas introduced into the compressor. Both theinlet and outlet openings respectively merge into supply and dischargemanifolds 108 and 110. A plate 112 of ringlike configuration is held inposition between the cylinder head and the housing by the bolts 100 andits primary function is to serve as a seat for the valves controllingthe flow of gas into and from the compression chamber.

As more specifically shown in FIGS. 9, 10 and 11, the plate 112 isequipped with gas supply openings 114 and gas discharge openings 116,the two sets of openings being isolated from one another by a centrallydisposed member 118. Although many different designs of valvingarrangements are available and may be employed for controlling thesupply and discharge openings, the inlet valve shown consists of a fiatleaf 120 having multiple cantilever reeds 121 which elastically deflectto an open position, as shown in FIG. 11, to admit gas when cylinderpressure is lower than inlet manifold pressure. Likewise, the outletvalve 122 is similar and deflects outward to exhaust gas into thedischarge manifold when cylinder pressure exceeds discharge manifoldpressure. An end of each valve is located in a depression 130 formed inthe compressor housing and is immovably fixed therein by the plate 112.The dotted lines in FIG. 11 illustrate the position of both valves whenin an open position. Valve stop 126 limits valve fiexure to apredetermined amount.

By utilizing valves of the above design having reeds which elasticallydeflect, the 'flow of gas through the supply and discharge openings inthe plate can be controlled very effec-- tively. The dimensions of eachreed are slightly larger than the openings in the plate with which theyalign to assure complete control over the gas flow to and from thecompressor.

In operation, when the solenoid coil 36 is energized, theelectromagnetic forces produced moves the armature and the attachedpiston in a direction to compress gas in the cylinder in the same manneras described in relation to FIGS. l4. The spring coils simultaneouslycompress and as the piston reaches the outer end of its stroke, thepressure in the compression chamber overcomes the valve 122 biasingforces plus the prevailing pressure in the outlet, and the dischargevalve 122 then opens to permit the discharge of gas through the openings116 prior to flowing through the manifold 110 and the outlet. The springreaction forces, coupled with the electromagnetic forces acting on thearmature then take effect, and cause the piston to travel on its returnstroke. In so doing, the pressure in the compression chamber drops andpermits the reeds 121 of the inlet valve 118 to deflect to an openingposition, thus drawing in a fresh supply ofgas for compression.Throughout the above described operation, the spring coils act to keepthe piston moving parallel to the cylinder walls, thus eliminating theneed to provide liquid lubrication to the compressor components. Thepiston-cylinder clearance space is designed to a close tolerance, but itis obvious that gas from the compression chamber will leak by the pistonas it reciprocates on the cylinder axis during operation. If necessary,the leakage gas can be used to provide some gas lubrication of thepiston by incorporating special profiles, such as tapers, on thecylindrical surface of the piston and/or the cylinder.

As described in connection with the embodiment of FIG. 1, themass-flexure system comprising the piston, armature and the springs ismade to have a natural frequency of vibration which corresponds with thefrequency of the electromagnetic forces generated by the voltage appliedto the coil to provide a resonant mode of vibration which permits thecompressor to operate at its highest efficiency and with the maximumgeneration of power in the system.

The major benefits or advantages accruing from complete elimination ofbearings in the compressor parts include removal of the need forlubrication of any kind in the piston actuation system. Since lubricantsare not required, gases can be pumped without concern for contaminationwhich otherwise would occur in compressors lubricated with oil or otherlubricating mediums. The complete elimination of all wearing orsurface-to-surface contacting of parts not only reduces the number ofcomponents needed, but also provides a degree of reliability notattainable in compressors of similar design. The

mechanical efficiency likewise is increased because the only losses arethose resulting from material damping in the resonant spring andexternal windage effects, both of which are very small. Anotherimportant advantage is the pressure-flow characteristics of thecompressor which can be adjusted almost instantly by simply changing thevoltage applied to the drive solenoid to obtain variance in the pistonstroke. It is apparent that this capability permits use of thecompressor in a wide variety of applications and in any particularinstallation where it is desired to maintain maximum compressionefficiency over a wide range of operating flows and pressures.

The modification illustrated in FIGS. 12 and 13 is substantiallydifferent from that previously described, although both designs havemany common features. The major difference lies in the selection andlocation of the springs used in the mass-flexure system.

As indicated previously, it is essential that the spring coils in thecompressor be disposed in a manner to obtain movement of the piston onthe center line of the cylinder and thereby not run or otherwise contactthe cylinder walls. Although the previous modifications included inconstruction wherein the coils of the spring were nested in one another,the arrangements shown in FIGS. 12 and 13 utilize a multiplicity ofsprings which are circumferentially disposed inside the piston.

Referring more specifically to FIGS. 12 and 13, it will be seen that thecompressor includes a solenoid 34 comprising a coil 36 wound on alaminated magnetic core 38, and an armature which is designed for axialmovement in the solenoid when the coil is energized from a power source.The housing 20 for the compressor serves to form a cylinder 22 having acylinder head 24 including a discharge valve 26 of the type discussedand illustrated in FIGS. 9-11. The housing includes inlet ports 72disposed around the housing for introducing a gas directly into thepiston-cylinder wall clearance space for lubrication purposes in thesame manner as that illustrated in FIG. 1.

The mass-flexure system of the compressor includes a connecting rodinterconnecting the armature 40 with the piston 28. A nut 152 securesthese parts together and the cylinder head 24 is designed with a recesshaving a configuration complementary to the nut 152' for permitting thepiston to utilize the full compression area in the compression space.Obviously, the bolt may terminate in a depression formed in the pistonhead should that design be more desirable. The connecting rod 150 isequipped with a plate 154 and is arranged to have a multiplicity ofsprings 156 located on opposite sides of the plate as shown in bothFIGS. 12 and 13. One end of each of the springs is fitted around acollar 158 formed on opposite sides of the plate while the other end ofthe springs fit over similar collars 160, respectively disposed on abase plate 162 and on the inner surface of inner housing 164. The innerhousing 164 is immovably positioned within the piston and is equippedwith a projecting flange 166 located between the housing 20 and plate162 and firmly anchored thereto by bolts, not shown.

In operation, it will be seen that as the coil 36 is energized, theelectromagnetic forces produced cause the armature 40, connecting rod150 and piston 28 to reciprocate in cylinder 22. In doing so, gas isintroduced through inlet ports 56 and upon upward movement of the pistonin the cylinder, gas is compressed and when it reaches a value greaterthan the pressure in the manifold holding the discharge valve 26 closed,the valve reeds lift and gas is discharged from the cylinder into themanifold for distribution to an attached system. As the piston travelson its return stroke, the valve 26 will close as the manifold pressureexceeds that in the compression space and further movement of the pistoneventually uncovers ports 56 for the introduction of a fresh supply ofgas for compression.

As in the previous modifications, the mass-flexure system comprising thearmature 40, connecting rod 150, piston 28, and springs 156, operatesin'resonance with the electromagnetic forces produced by the coil duringalternating cycles. When the resonant condition is reached, thecompressor operates at maximum efficiency and power in supplyingcompressed gas to the attached system. In view of uniform distributionof springs 1156 on both sides of plate 154, the piston 28 will moveaxially on the cylinder center line and not engage or rub the cylinderwalls. As in the previous modifications, gas may be introduced into thepiston-cylinder wall clearance space to provide a hydrostatic lubricantand help assure that the piston will not contact the cylinder walls.

It will be evident that on the compression stroke, the springs 156 onthe upper side of plate 154 will compress while those below the platewill expand from their initial precompressed condition. Although thisdesign of compressor utilizes a multiplicity of springs, the totaleffect produced is one of providing great reliability in maintainingmovement of the piston one the cylinder axis. The construction shown inFIGS. 12 and 113 lends itself to some of the variations in thecompressor structure heretofor discussed in this application. Forexample, the cylinder head may be equipped with both an inlet valve anddischarge valve for controlling the admission and discharge of gas fromthe compression area. I

The modification illustrated in FIGS. 14-16 is similar in operation tothat previously described but differs in the changes made to acquire amore compact structure and one that is more economical to manufacture.Such differences include the means for connecting the armature to thepiston and spring supporting yoke and in providing a design of yoke andpiston to provide a compact structure necessary for decreasing the sizeof the compressor. As shown, the modification in cludes a housing 22,cylinder 26 and piston 20 adapted for reciprocation therein. A valvedinlet 132. and outlet 134 are provided for controlling the supply of gasto the compressor and for discharging it after compression into a closedsystem. One side of the housing is equipped with outwardly extendingprojections which serve as a base for supporting arms 136, 138 and Mwhich terminate in a member designed for attachment to the iron core 38of the solenoid 34.

Referring more specifically to the parts, the piston is of essentiallythe same design as the modifications previously described and can bemade to provide a relatively close clearance between the piston andcylinder wall for piston sealing purposes. The ratio of radial clearanceto piston diameter has typically been about 0.0005 in those designswhere gas is used for lubricating the piston. The relative motion of thepiston with respect to the cylinder wall in such designs is used togenerate a self-acting (i.e. hydrodynamic) fluid film bearing action tocenter the piston in accordance with well known gas bearing principles.Alternatively, instead of utilizing the relatively moving parts forgenerating the pressure needed for lubrication purposes, a small bleedflow can be taken from the compressor discharge to pressurize the spacebetween the piston and cylinder walls as described in relation to FIGS.llil, thus providing a hydrostatic bearing centering action. In eithercase, the actual gas which is being pumped is the gas which is used tolubricate the piston. The close clearance between piston and cylinderreduces piston bypass leakage to small values, while the fluid-filmcaring action eliminates rubbing contact between the piston and cylinderwalls. As a result a very long-lived, contamination-free piston sealingmeans is obtained.

In one particular compressor having a piston radialclearance-to-strokeratio of 0.005, the fluid-film bearing function was completelyeliminated. In this arrangement, the piston was held centered within thecylinder clearance space solely by the spring. Performance of thiscompressor showed that after more than 4,000 hours of compressoroperation, no evidence of performance deterioration or piston wear isdetectable. The valves I and 122 controlling the inlet and outlet arethe same as that described in relation to FIGS. 9, 10, Ill and include aplate having openings controlled by flexible fingers attached to thebody of each valve.

As illustrated, the compressor is contained within a hermetically sealedenclosure 142, the purpose being to provide a system for isolatingvibrations during compressor operation and for permitting the use of aconvenient arrangement for introducing gas to the compressor and fordischarging it into a reservoir before delivery to the external system.To facilitate the entry of gas, a fitting M4 is secured in fluid tightrelationship with the enclosure and a flexible tube 146 interconnectsthe fitting with a racket M7 attached to the housing by bolts 14% whichserves as a gas inlet to the compressor. Since the compressor vibratesduring the course of performing its compression function, the flexibletube M6 effectively serves as a vibration isolating device, thusminimizing the magnitude of the mechanical vibrations which otherwisewould be transmitted to the enclosure M2.

The yoke arrangement consisting of arms I36, I38 and M0 serve the dualfunction of supporting solenoid 34 and forming a base for anchoring theends of the spring 32 in a firm position. To provide an optimumsupporting arrangement, the three arms are located on one side of thecompressor and are held to the housing by means of bolts M9. As seen inFIG. 16, the arms appear only on one side but may be arranged to extendaround the compressor if desired. The other ends of the arms terminatein a supporting arrangement which immovably holds the iron core 36 onthe axis of the cylinder. As previously described, the spring element 32consists ofindependent springs 1150, 152, 154 each respectively beingsecured at one end by a clamp 156 or similar securing means. The upperends of each of the springs are attached to the housing 22 by a similararrangement of clamps.

Although resonant compressor designs have been proposed which yield atheoretical balance of the dynamic forces acting on the frame of thecompressor, such designs are not generally attractive from a productioncost standpoint. In all of the modifications discussed herein, thedynamic frame forces are unbalanced. It is necessary, therefore, tovibration-isolate the compressor from the base structure. To this end,and to obtain effective, yet inexpensive, isolation, the compressor issuspended from the inside of the hermetically sealed enclosure M2. Thesupporting means comprises a bar 162 welded or otherwise affixed to theinside surface of the enclosure. A pair of springs I64 interconnect thebar with a pair of spaced brackets 166, which as shown in FIG. 16,extend across the compressor with their ends terminating in brackets 168or similar securing means. The vibratory forces generated in thecompressor during operation are transmitted through the brackets to thesprings 1164 and bar 1162 where they are ab sorbed and dissipated by thesprings and the enclosure 142.

In operation, gas drawn into the inlet is delivered to the compressionarea through the reed-type valves I20, and after compression, the valves122 are forced open and gas is delivered to the cavity formed by theenclosure 142. The cavity forms the function of a reservoir and deliversthe compressed gas through an outlet 270 located in a side of theenclosure. The compressor operates in the same manner as in theembodiments of FIGS. 5-11. As it vibrates during operation, the springsI68 are caused to expand and contract in response to the vibrationsproduced by the moving parts. It will be apparent that the compressorcan be operated in any attitude simply by providing a multiplicity ofsprings 163 around the various parts of the compressor.

As disclosed herein, the approach to piston sealing and for preventingthe piston from rubbing the cylinder walls, is to provide a very closeclearance between the piston and cylinder wall in those designs where aportion of the process gas is used for piston lubrication purposes. Therelative motion of the piston with respect to the cylinder wall is usedfor generating a self-acting, i.e. hydrodynamic, a gas-bearing action tocenter the piston; or a small bleed flow can be taken from thecompressor discharge to pressurize the space between the piston andcylinder walls to provide a hydrostatic gas bearing centering action tocenter the piston. To insure minimum manufacturing costs, thepiston-cylinder clearance is increased, while still maintainingacceptable leakage, and the piston is kept centered in the cylinder bythe sole action of the spring connected to the housing and the piston.In either case, a very low-friction, long-life, contamination-free,piston sealing compressor is obtained. In view of the above, it will beapparent that many modifications and variations are possible in light ofthe above teachings. It therefore is to be understood that within thescope of the appended claims, the invention may be practiced other thanas specifically described.

What we claim as new and desire to secure by United States LettersPatent is:

We claim:

1. A resonant piston pump comprising:

a housing having a cylinder and a piston adapted for reciprocatingmovement therein;

a fluid inlet and outlet in said cylinder for admitting fluid throughthe inlet to be operated on by said piston, and after such operation,discharging said fluid through the outlet;

means for controlling the flow of fluid through said inlet and outlet;

coiled spring means anchored at one end to said housing and connected atthe other end to said piston, the arrangement being such that saidspring means in undergoing compression and expansion as a result ofpiston movement during operation, causes the piston to movesubstantially parallel with and out of contact with the walls of saidcylinder; and

drive means including a movable member connected with said piston fordriving said piston, springs and the movable member in a resonant modeof vibration to thereby achieve the maximum transfer of power andefficiency in said pump.

2. The combination according to claim 1 wherein said piston comprises ahead and hollow skirt integrally attached thereto;

a shaft interconnecting said movable member with the piston head forimparting a reciprocal motion to the piston when the drive means isactuated; and

wherein said spring means are positioned within said hollow piston.

3. Combination according to claim 2 wherein said spring means ismachined from solid stock material;

said spring means comprising a plurality of separate coils each of whichconstitutes a spring integrally merged into said piston head atone endand merged into an integrally formed flangelike member at their otherends; and

means securing said flangelike member to said housing for anchoring saidspring means firmly in position.

4. The combination according to claim 3 wherein said inlet comprises amultiplicity of ports spaced around the cylinder peripherally; and

wherein said inlet is controlled by a valve positioned in the cylinderhead.

5. The combination according to claim 4 wherein said valve consists of aflexible plate having a plurality of flexible reeds which deflect to anopen position when the pressure in said compression space exceeds apredetermined value, and closes when the pressure in said compressionspace decreases below a predetermined valve.

6. The combination according to claim 5 wherein said outlet is connectedthrough passages in said housing for discharging fluid into thepiston-cylinder clearance space for lubricating said piston duringoperation.

7. The combination according to claim 5 wherein said drive meanscomprises a solenoid and wherein said movable member comprises aarmature; and

means interconnecting said shaft with said armature.

8. A resonant piston pump comprising:

a housing having a cylinder and a piston designed for reciprocationtherein;

a head closing the upper end of said cylinder to provide an operationarea with said piston during operation;

a valved inlet and outlet in said head for controlling the admission anddischarge of a fluid to and from said operation area; drive meanssupported by said housing and including a movable member connected withsaid piston to cause it to reciprocate in said cylinder;

spring means comprising a plurality of independent coils each having oneend immovably attached to said housing and their other ends to anextension on said movable member, said coils being positioned andarranged in a manner to uniformly transmit or absorb forces to or fromthe piston respectively, to thereby minimize piston lateral thrustforces and cause the piston to move substantially parallel to thecylinder walls;

said piston, movable member and spring means comprising a mass-flexuresystem designed to a natural vibration frequency corresponding to thedriving force frequency of said drive means to drive said pump in aresonant mode of vibration.

9. The combination according to claim 8 wherein said extension on saidmovable member comprises a yoke having outwardly extending projections;and means securing an end of each of said coils respectively to thecorresponding projections on said yoke.

10. The combination according to claim 9 wherein said coils are disposedaround and are of a greater diameter then said cylinder.

11. The combination according to claim 9 wherein said drive meanscomprises iron core having an electrical coil wound thereon; and

wherein said movable member comprises an armature, so

that upon energization of said coil, the electromagnetic forces producedcauses said mass-flexure system to move in resonance therewith.

12. The combination according to claim 11 wherein said armature and theiron core opening into which it reciprocates have complementary surfacesof substantially square configuration.

13. The combination according to claim 11 wherein said valved inlet andoutlet comprises openings formed in said cylinder head and valvescontrolling fluid flow therethrough;

each of said valves comprising a flat element equipped with multipleflexible reeds designed to yield in response to the pressure in saidcompression area for controlling fluid flow through the inlet andoutlet; and

a stop in said cylinder head for limiting the amount said reeds candeflect when moved to an open position.

M. The combination according to claim 13 wherein said valves are mountedon a plate secured between said cylinder head and housing.

UNITED STATES PA"EN'I ()FFE'CE CERTIFICATE OF CORRECTION patent 9 DatedJune 28, 1971 inventor) Peter W. (Iurwen wt 11 I "W It is certified thaterror appears in the above-identified patent and that said LettersPatent are hereby corrected as shown below:

line 56 Column 9, line 14, "one" should read on "earing" should readbearing Column 10, line 6, "racket" should read bracket line 51 "270"should read 170 "predetermined valve" should read Column 11, line 58,predetermined value Signed and sealed this 2nd day of May 1972.

(SEAL) Attest:

ROBERT GOTTSCHALK EDWARD M.FLETCHER,JR.

Commissioner of Patents Attesting Officer FORM F G-W50 (10-69 USCOMM-DCwave-ps9 9 U S GOVERNMENY PRINTING OFFICE: 1969 0-356-331

